Variable capacity screw compressor and method

ABSTRACT

A variable capacity screw compressor comprises a suction port, at least two screw rotors and a discharge port being configured in relation to a selected rotational speed that operates at least one screw rotor at an optimum peripheral velocity that is independent of a peripheral velocity of the at least one screw rotor at a synchronous motor rotational speed for a rated screw compressor capacity. A motor is configured to drive the at least one screw rotor at a rotational speed at a full-load capacity that is substantially greater than the synchronous motor rotational speed at the rated screw compressor capacity. A variable speed drive receives a command signal from a controller and generates a control signal that drives the motor at the selected rotational speed. A method for sizing at least two variable capacity screw compressors and a refrigeration chiller incorporating a variable capacity screw compressor are separately presented.

CROSS-REFERENCE TO RELATED APPLICATIONS

Not applicable presently.

FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT

None.

BACKGROUND OF THE INVENTION

This invention relates generally to screw compressors for refrigerationsystems. Particularly, the invention relates to screw compressors forrefrigeration systems that are operable at a rotational speedsubstantially above a synchronous motor rotational speed for a ratedcapacity and that have an inlet port, a discharge port, screw rotors anda selected rotational speed configured such that the selected rotationalspeed drives at least one screw rotor at an optimum peripheral velocityindependent of the rated capacity of the screw compressor. Configuringthe inlet port, the discharge port and the screw rotors together withthe selected rotational speed allows for a plurality of screwcompressors to be produced with different rated capacities, such thateach screw compressor delivers approximately the same high efficiencyfrom operation of at least one screw rotor at the same optimumperipheral velocity.

Compressors in refrigeration systems raise the pressure of a refrigerantfrom an evaporator pressure to a condenser pressure. The evaporatorpressure is sometimes referred to as the suction pressure. The condenserpressure is sometimes referred to as the discharge pressure. At thesuction pressure, the refrigerant is capable of cooling a desiredmedium.

Many compressor types, including rotary screw compressors, are used insuch refrigeration systems. Rotary screw compressors, or screwcompressors, are positive displacement, volume reduction devices.

Screw compressors typically employ a male screw rotor and female screwrotor, sometimes in the form of helical intermeshing rotors. A screwcompressor arrangement having an intermeshing male screw rotor andfemale screw rotor is also sometimes called twin screw compressor.Helical intermeshing rotors have a geometry or profile that is defined,in part, by the number of rotor lobes, the wrap angle, the length of therotors and the diameter of the rotors, for example. One profile ofrotors is not universal to all screw compressors.

The intermeshing screw rotors are mounted for rotation in a workingchamber defined by a compressor or rotor housing. The working chamberconsists of a volume shaped as a pair of parallel intersectingflat-ended cylinders and is closely toleranced to the exteriordimensions and shapes of the intermeshing screw rotors. The clearancetolerances between the surface of the working chamber and exterior ofthe intermeshed screw rotors can be as low as a few micrometers.

The screw compressor has a low pressure end and a high pressure end. Thelow pressure end contains a suction port. The high pressure end containsa discharge port. The low pressure end and the high pressure end eachopen into the working chamber.

In conventional operation of refrigeration-based systems, thecounter-rotation of the intermeshing screw rotors draws a mass ofrefrigerant gas at suction pressure into the suction port from a suctionarea at the low pressure end of the compressor. The refrigerant isdelivered through the suction port to a compression pocket having achevron shape, sometimes called a flute space. The compression pocket isdefined by the intermeshed rotors and the interior wall of the workingchamber.

As the intermeshing screw rotors rotate, the compression pocket isclosed off from the suction port. Gas compression occurs as thecompression pocket volume decreases as the intermeshing screw rotorsrotate. The compression pocket is circumferentially and axiallydisplaced to the high pressure end of the compressor by the rotation ofthe intermeshing screw rotors and comes into communication with thedischarge port. The compressed refrigerant gas is discharged through thedischarge port from the working chamber.

It is often desirable to operate such screw compressors at part-loadconditions (i.e. when full capacity operation is not required). Toimprove performance at part-load conditions, several approaches havebeen employed.

One approach that has been employed is the use of slide valvearrangements. These slide valve arrangements may take the form of aslide valve assembly. A valve portion of a slide valve assembly istypically disposed within the working chamber. The slide valve is opened(or closed) to varying degrees to vary capacity from a full-load to apart-load. When the slide valve is opened (or closed) to vary capacity,the working chamber and the screw rotors are exposed to a larger (orlesser) extent to the suction pressure. Generally, the slide valveassembly accommodates part-load performance, because the working chamberand the screw rotors that are exposed to the suction pressure cannotengage in the compression process and thereby, the screw compressor'scapacity is reduced proportionately. Improved part-load performance canbe achieved with these slide valve arrangements through control of thedischarge port.

Inclusion of such slide valve arrangements increases the mechanicalcomplexities of the screw compressor because of the increase in thephysical part count, the precision machining needed for sealing, andtheir reduction of efficiency at part load by, for example, furtherincreasing leakage area along the slide valve interface with the rotors.An efficiency penalty also can result from inclusion of slide valvestypically because of the additional clearance from the rotors that isrequired. Due, in part, to the increase of the mechanical complexitiesassociated with employing slide valve arrangements, slide valves alsocan contribute to manufacturability difficulties and to potentialreduced reliability in day-to-day operation.

Another approach has been to employ poppets. Poppets also can serve as amechanical unloader. Use of poppets in screw compressors suffer fromsimilar drawbacks as slide valves. For example, poppets requireadditional manufacturing complexities due to increased part count andmachining.

Yet another approach that has been employed to improve part-loadperformance is the use of variable speed drives (VSDs). VSDs controlmotor loading by varying the speed that a motor drives the intermeshingscrew rotors. VSDs typically vary the frequency and/or voltage providedto the motor. This frequency or voltage variance can allow the motor toprovide a variable output speed and power in response to the load on themotor.

Employing VSDs in conventional screw compressors can cause three to sixpercent or more loss in efficiency at full-load capacity. At the sametime, VSDs have a cost on the order of a screw compressor. Anotherchallenge with employing VSDs to operate conventional motors is thatthey reach their peak efficiency at their rated speed. As a result,motor efficiency drops at lower speeds. Such reduced theoreticalperformance compromises the energy savings level at part-loadconditions. In an attempt to minimize the performance compromise ofVSDs, gear train arrangements have been employed to optimize tip speedand rotational input into the screw compressor. Employing gear trainspresent challenges of their own, including gear train-related parasiticlosses, added lubrication, added maintenance, more noise and largerspace requirements.

Regardless of which approach is employed to achieve part-loadperformance, neither slide valve arrangements nor variable speed drivesused in conventional screw compressors have achieved compact, variablecapacity screw compressors that are operable at a rotational speedsubstantially above a synchronous motor rotational speed for a ratedcapacity of the screw compressor, where the inlet port, the dischargeport, screw rotors and the selected rotational speed are configured suchthat the selected rotational speed drives at least one screw rotor at anoptimum peripheral velocity independent of the rated capacity of thescrew compressor. Nor have the conventional screw compressors delivereda variable capacity screw compressor configuration where the inlet port,the discharge port and the screw rotors are sized together with theselected rotational speed such that a plurality of screw compressorswith different rated capacities can be produced, where each screwcompressor delivers approximately the same high efficiency fromoperation of at least one screw rotor at the same optimum peripheralvelocity.

BRIEF SUMMARY OF THE INVENTION

According to an embodiment of the present invention, a variable capacityscrew compressor comprises a rotor housing, a motor, and a variablespeed drive. The rotor housing comprises a suction port, a workingchamber, a discharge port, and at least two screw rotors that comprise afemale screw rotor and a male screw rotor being positioned within theworking chamber for cooperatively compressing a fluid. The suction port,the at least two screw rotors and the discharge port are configured inrelation to a selected rotational speed. The selected rotational speedoperates at least one screw rotor at an optimum peripheral velocity thatis independent of a peripheral velocity of the at least one screw rotorat a synchronous motor rotational speed for a rated screw compressorcapacity. A motor is operable to drive the at least one screw rotor at arotational speed at a full-load capacity that is substantially greaterthan the synchronous motor rotational speed at the rated screwcompressor capacity. A variable speed drive receives a command signalfrom a controller and generates a control signal that drives the motorat the rotational speed.

In another embodiment, a method for sizing at least two screwcompressors is provided. The target capacity for each screw compressoris selected. Each screw compressor has a different rated capacity andfurther comprises a suction port, a working chamber, a discharge port,and at least two screw rotors being positioned within the workingchamber for cooperatively compressing a fluid. The rotational speed isselected to operate at least one screw rotor in each screw compressor atan approximately constant optimum peripheral velocity that isindependent of the rated capacity of each screw compressor. The suctionport, the at least two screw rotors and the discharge port areconfigured together with the rotational speed for each screw compressor.

In another embodiment, a refrigeration chiller, having at least onerefrigeration circuit, comprises a variable capacity screw compressor,condenser, expansion valve and evaporator. The variable capacitycompressor comprises a rotor housing, a motor housing and a variablespeed drive. The rotor housing further comprises a suction port, aworking chamber, a discharge port, and at least two screw rotors thatcomprise a female screw rotor and a male screw rotor being positionedwithin the working chamber for cooperatively compressing a fluid. Thesuction port, the at least two screw rotors and the discharge port areconfigured in relation to a selected rotational speed. The selectedrotational speed provides at least one screw rotor to operate at anoptimum peripheral velocity that is independent of a peripheral velocityof the at least one screw rotor at a synchronous motor rotational speedfor a rated screw compressor capacity. The motor housing furthercomprises a motor, the motor is operable to drive the at least one screwrotor at a rotational speed at a full-load capacity that issubstantially greater than the synchronous motor rotational speed at therated screw compressor capacity. The variable speed drive is configuredto receive a command signal from a controller and to generate a controlsignal that drives the motor at the rotational speed. A condenser iscoupled to the discharge port of the variable capacity screw compressor.The condenser is configured to cool and condense fluid received from thedischarge port. An expansion valve is coupled to the condenser. Theexpansion valve is configured to evaporate at least a portion of fluidreceived from the condenser by lowering pressure of fluid received fromthe condenser. An evaporator is coupled to the expansion valve. Theevaporator is configured to evaporate fluid received from the expansionvalve and to provide fluid to the suction port of the variable capacityscrew compressor.

Those skilled in the art will appreciate advantages and superiorfeatures of the above embodiments, together with other important aspectsthereof upon reading the detailed description which follows inconjunction with the drawings. Additional advantages and features of theinvention will become more apparent from the description of anembodiment of the present invention and the claims which follow.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

The following figures include like numerals indicating like featureswhere possible:

FIG. 1 illustrates an embodiment of the present invention thatincorporates a screw compressor arranged as part of a refrigerationchiller system.

FIG. 2 illustrates a cross sectional view of a screw compressoraccording to an embodiment of the present invention.

FIG. 3 illustrates an additional cross sectional view of a screwcompressor according to an embodiment of the present invention.

FIG. 4 illustrates an embodiment of a refrigeration chiller andcontroller system according to an embodiment of the present invention.

DETAILED DESCRIPTION OF AN EMBODIMENT

As a preface to the detailed description, as used in this specificationand the appended claims, the singular forms “a,” “an,” and “the” alsoinclude plural referents, unless the context clearly dictates otherwise.References in this specification to “one embodiment,” “an embodiment,”“an example embodiment,” etc., indicate that the described embodimentmay include a particular feature, structure, or characteristic; however,every embodiment may not necessarily include the particular feature,structure, or characteristic. When a particular feature, structure, orcharacteristic is described in connection with an embodiment, otherembodiments may incorporate or otherwise implement such feature,structure, or characteristic whether or not explicitly described.

Referring now to FIGS. 1-4, components of a chiller or chiller system 10are illustrated. Chiller 10 includes many other conventional featuresnot depicted for simplicity of the drawings.

Chiller system 10 is directed to refrigeration systems. Chiller 10 is inthe range of about 20 to 500 tons or larger, particularly where therefrigeration system includes a multiple stage compressor arrangement.Persons of ordinary skill in this art will readily understand thatembodiments and features of this invention are contemplated to includeand apply to, not only single stage compressors/chillers, but also to(i) multiple stage compressors/chillers and (ii) single and/ormultistage compressor/chillers operated in parallel.

As shown, chiller 10 comprises a screw compressor system 12 (alsosometimes referred to as a screw compressor 12), a condenser 14, and anevaporator 20, all of which are serially connected to form a semi- orfully-hermetic, closed-loop refrigeration system. Chiller 10 maycirculate a fluid 80 (such as, for example, a refrigerant) to controlthe temperature in a space such as a room, home, or building. The fluid80 may be circulated to absorb and remove heat from the space and maysubsequently reject the heat elsewhere.

Fluid 80 may be a refrigerant. The refrigerant may be selected from anazeotrope, a zeotrope or a mixture or blend thereof in gas, liquid ormultiple phases. For example, such refrigerants may be selected from:R-123, R-134a, R-1234yf, R-410A, R-22 or R-32. Because embodiments ofthe present invention are not restricted to the refrigerant chosen,embodiments of the present invention are also adaptable to a widevariety of refrigerants that are emerging, such as low global warmingpotential (low-GWP) refrigerants.

FIG. 1 illustrates the condenser 14. Condenser 14 is shown as a shelland tube flooded-type. The condenser 14 can be arranged as a singleevaporator or multiple evaporators in series or parallel, e.g.connecting a separate or multiple evaporators to each compressor.Condenser 14 may include condenser tubing 16. Fluid 80 may pass acrossthe condenser tubing 16 through which cool air or cool liquid flows.

Condenser 14 may be fabricated from carbon steel and/or other suitablematerial, including copper alloy heat transfer tubing. Condenser tubing16 can be of various diameters and thicknesses, and comprised typicallyof copper alloy. In addition, condenser tubing 16 may be replaceable,mechanically expanded into tube sheets and externally finned seamlesstubing. Other known types of condenser 14 are contemplated.

Condenser 14 may be configured to communicate fluid 80 from a dischargepassage 36. Discharge passage 36 may be configured to receive the fluid80, or may be coupled to the condenser 14 through an oil separator 24,as depicted in FIG. 1. Other configurations are contemplated. The oilseparator 24, when employed, separates oil from the fluid 80 and returnsthe oil via an oil supply passage 26 to the screw compressor 12 forreuse. The oil may be reused to, for example, cool the fluid 80, coolscrew rotors 42, seal the interfaces between the screw rotors 42themselves, seal the interfaces between the screw rotors 42 and thewalls of a working chamber 44, and/or lubricate bearings 46, 48.

Condenser 14 may transform the fluid 80 from a superheated vapor to asaturated liquid. As a result of the cool air or cool liquid passingacross the condenser tubing 16, fluid 80 may reject or otherwise deliverheat from the chiller 10 to another fluid, like air or liquid, in a heattransfer relation, which in turn carries the heat out of the system.

An expansion valve 18 may employed, as shown in FIG. 1. Expansion valve18 may be configured to receive fluid 80 from condenser 14. Fluid 80received from condenser 14 typically is in a thermodynamic state knownas a saturated liquid. The expansion valve 18 may abruptly reduce thepressure of the fluid 80. The abrupt pressure reduction may causeadiabatic flash evaporation of at least a portion of the fluid 80. Inparticular, the adiabatic flash evaporation may result in a liquid andvapor mixture of the fluid 80 that has a temperature that is colder thanthe temperature of the space to be cooled.

Evaporator 20 is shown in FIG. 1 as a shell and tube flooded-type. Theevaporator 20 can be arranged as a single evaporator or multipleevaporators in series or parallel, e.g. connecting a separate ormultiple evaporators to each compressor. Evaporator 20 may includeevaporator tubing 22. Fluid 80 may pass across the evaporator tubing 22through which cool air or cool liquid flows.

Evaporator 20 may be fabricated from carbon steel and/or other suitablematerial, including copper alloy heat transfer tubing. Evaporator tubing22 can be of various diameters and thicknesses, and comprised typicallyof copper alloy. In addition, evaporator tubing 22 may be replaceable,mechanically expanded into tube sheets and externally finned seamlesstubing. Other known types of evaporator 20 are contemplated.

Evaporator 20 is configured, as illustrated in FIG. 1, to receive fluid80 communicated from the expansion valve 18. Fluid 80 received by theevaporator 20 in the refrigeration loop may be relatively colder than itwas when discharged from the screw compressor 12. The oil returnapparatus 28, when employed, separates oil from the fluid 80 and returnsthe oil via an oil return passage 30 to the screw compressor 12 forreuse. The oil may be reused to, for example, cool the fluid 80, coolscrew rotors 42, seal the interfaces between the screw rotors 42themselves, seal the interfaces between the screw rotors 42 and thewalls of a working chamber 44, and/or lubricate the bearings 46, 48.

The evaporator 20 may absorb and remove heat from the space to becooled, and the condenser 14 may subsequently reject the absorbed heatto air or liquid that carries the heat away from the space to be cooled.In operation, warm air or liquid may be circulated from the space to becooled across the evaporator tubing 22. The warm air or liquid passingacross the evaporator tubing 22 may cause a liquid portion of the coldfluid 80 to evaporate. At the same time, the warm air or liquid passedacross the evaporator tubing 22 may be cooled by the fluid 80. It shouldbe understood that any configuration of the condenser 14 and/orevaporator 20 may be employed that accomplishes the necessary phasechanges of fluid 80.

The chilled or heated water is pumped from the evaporator 20 to an airhandling unit (not shown). Air from the space that is being temperatureconditioned is drawn across coils in the air handling unit thatcontains, in the case of air conditioning, chilled water. The drawn-inair is cooled. The cool air is then forced through the air conditionedspace, which cools the space.

Additionally, though not shown, an economizer 32 may be incorporated toinclude an economizer cycle. Economizer 32 or a subcooling cycle (notshown), or both, may be employed in the refrigeration cycle and returnthe fluid 80 to the screw compressor 12 via suction passage 34 or otherpassage (not shown) depending on the configuration required theapplication conditions.

Referring to FIGS. 2 and 3, screw compressor 12 typically comprises arotor housing 40 and an electric motor housing 50. Screw compressor 12may be formed, all or in part, of gray cast iron, for example. Othermaterials may be used to form the screw compressor 12. Screw compressor12, according to embodiments of the present invention, facilitateshighly efficient operation at full-load and part-load conditions over apreselected screw capacity range.

Motor housing 50 houses a motor 52 in an embodiment of the presentinvention. Electric motor 52 may coupled to a variable frequency drive38. The electric motor 52 drives meshed screw rotors 42. Motor housing50 may be integral to the rotor housing 40.

The rotor housing 40 may have a low pressure end and a high pressure endthat each contain a suction port 76 and discharge port 78, respectively.Suction port 76 and discharge port 78 are in open-flow communicationwith the working chamber 44. The suction port 76 and the discharge port78 may each be an axial, a radial or a mixed (a combination of a radialand an axial) port.

The suction port 76 may receive the fluid 80 at a suction pressure and asuction temperature. The suction port 76 may receive fluid 80 fromsuction passage 34 in thermodynamic states known as a saturated vapor ora superheated vapor. The screw compressor 12 may compress the fluid 80as the screw compressor 12 communicates the fluid 80 from the suctionport 76 to the discharge port 78. Fluid 80 passing through the dischargeport 78 discharges into discharge passage 36.

Compressing the fluid 80 may also result in the fluid 80 beingdischarged at a discharge temperature that is higher than the suctiontemperature. The fluid 80 discharged from the discharge port 78 may bein a thermodynamic state known as a superheated vapor. Accordingly,fluid 80 discharged from the screw compressor 12 may be at a temperatureand a pressure at which the fluid 80 may be readily condensed with acooling air or a cooling liquid.

Suction port 76 and discharge port 78 are configured to minimize flowlosses, when at least one of the rotors 42 is operated at anapproximately constant peripheral velocity. The suction port 76 may belocated where fluid 80 exits the suction area of screw compressor 12 andis drawn into the working chamber 44. The suction port 76 may be sizedto be as large as possible to minimize, at least, the approach velocityof the fluid 80. The location of the suction port 76 in the rotorhousing 40 also may be configured to minimize turbulence of fluid 80prior to entry into the rotors 42.

Discharge port 78 may be sized larger than theoretically necessary toprovide a thermodynamic optimum size and thereby, reduce the velocity atwhich the fluid 80 exits the working chamber 44. The discharge port 78may be generally located where fluid 80 exits the working chamber 44 ofscrew compressor 12. The discharge port 78 location in the rotor housing40 may be configured such that the maximum discharge pressure can beattained in the rotors 42 prior to being delivered into the dischargepassage 36. In addition, screw compressor 12 may incorporate a muffler58 or other apparatus suitable for noise reduction.

Referring again to FIG. 3, rotors 42 are mounted for rotation in aworking chamber 44. The working chamber 44 comprises a volume that isshaped as a pair of parallel, intersecting flat-ended cylinders, and isclosely toleranced to the exterior dimensions and geometry of theintermeshed screw rotors 42. The plurality of meshed screw rotors 42 a,42 b may define one or more compression pockets between the screw rotors42 a, 42 b and the interior chamber walls of the rotor housing 40. Therotor housing 40 has little separation from the rotors 42. Milling,machine grinding or molding can be employed to achieve high accuracy andtight tolerances between rotors 42 flutes and lobes and the rotorhousing 40.

First screw rotor 42 a and second screw rotor 42 b are disposed in acounter-rotating, intermeshed relationship and cooperate to compress afluid. At least one of rotors 42 is cooperatively configured with motor52 to be operable at a rotational speed for a screw compressor capacitywithin a preselected screw compressor capacity range. The selectedrotational speed at full-load capacity is substantially greater than asynchronous motor rotational speed at a rated capacity (also referred toherein as rated screw compressor capacity) for screw compressor 12.

In the embodiment illustrated, rotor 42 a may be called a female screwrotor and comprise a female lobed/fluted body or working portion(typically a helical or spiral extending land and groove). Rotor 42 bmay be called a male screw rotor and comprise a male lobed/fluted bodyor working portion (typically a helical or spiral extending land andgroove).

Rotors 42 include shaft portions, which are, in turn, mounted to thehousing of screw compressor 12 by, for example, one or more bearings 46,48. The exemplary bearings 46, 48 will also be configured with tightclearances in relation to at least rotors 42 and rotor housing 40.

Compression of the fluid 80 in screw compressor 12 produces axial andradial forces. The configurations of embodiments of the presentinvention may also mitigate time varying and non-uniform rotor movementsand forces against chamber walls, bearings, and end surfaces of thescrew compressor 12 caused by the interaction of the screw rotors 42 a,42 b, the axial forces, and the radial forces.

As mentioned, a lubricating fluid, typically oil, may be delivered fromoil supply passage 26 or oil return passage 30 to the screw compressor12. The lubricating fluid provides cushioning films for the walls of theworking chamber 44, rotors 42 a, 42 b, and bearings 46, 48 of the screwcompressor 12, but does little to prevent the transmission of the timevarying and non-uniform axial and radial forces. The screw compressor 12may also utilize an expander (not shown), which may also be integral toscrew compressor 12, to recover energy available from the refrigerationcycle as the high pressure liquid expands through the expander to alower pressure.

The electric motor 52 in one exemplary embodiment may drive at least oneof the rotors 42 in response to command signals 62 received from thecontroller 60. The horsepower of preferred motor 52 can vary in therange of about 125 horsepower to about 2500 horsepower. Torque suppliedby the electric motor 52 may directly rotate at least one of the screwrotors 42. Employing motor 52 and variable speed drive 38, screwcompressor 12 of embodiments of the present invention may have a ratedscrew compressor capacity within the range of about 35-tons to about150-tons or more and have a full-load speed range within about 4,000revolutions per minute to about 15,000 revolutions per minute, when thefluid is an R-134a refrigerant.

While conventional types of motors, like induction motors, can be usedwith and will provide a benefit when employed with embodiments of thepresent invention, a preferred motor 52 comprises a direct drive,variable speed, hermetic, permanent magnet motor. Permanent magnet motor52 can increase system efficiencies over other motor types. The choiceof motor 52 may be affected by cost and performance considerations.

Referring to FIGS. 2 and 3, the permanent magnet motor 52 comprises amotor stator 54 and a motor rotor 56. Stator 54 consists of wire coilsformed around laminated steel poles, which convert variable speed drive38 applied currents into a rotating magnetic field. The stator 54 ismounted in a fixed position in the screw compressor 12 and surrounds themotor rotor 56, enveloping the rotor 56 with the rotating magneticfield. Motor rotor 56 is the rotating component of the motor 52 and mayconsist of a steel structure with permanent magnets, which provides amagnetic field that interacts with the rotating stator magnetic field toproduce rotor torque. In addition, permanent magnet motor 52 may beconfigured to receive variable frequency control signals and to drivethe at least two screw rotors per the received variable frequencycontrol signals.

The motor rotor 56 may have a plurality of magnets and may comprisemagnets buried within the rotor steel structure or be mounted at therotor steel structure surface. Motor rotor 56 surface mount magnets aresecured with a low loss filament, metal retaining sleeve or by othermeans to the rotor steel support. Further manufacturing, performance,and operating advantages and disadvantages can be realized with thenumber and placement of permanent magnets in the motor rotor 56. Forexample, surface mounted magnets can be used to realize greater motorefficiencies due to the absence of magnetic losses in interveningmaterial, ease of manufacture in the creation of precise magneticfields, and effective use of rotor fields to produce responsive rotortorque. Likewise, buried magnets can be used to realize a simplermanufactured assembly and to control the starting and operating rotortorque reactions to load variations.

The performance and size of the permanent magnet motor 52 is due in partto the use of high energy density permanent magnets. Permanent magnetsproduced using high energy density magnetic materials, typically atleast 20 MGOe (Mega Gauss Oersted), produce a strong, more intensemagnetic field than conventional materials. With a motor rotor 56 thathas a stronger magnetic field, greater torques can be produced, and theresulting motor 52 can produce a greater horsepower output per unitvolume than a conventional motor, including induction motors. By way ofcomparison, the torque per unit volume of permanent magnet motor 52 canbe at least about 75 percent higher than the torque per unit volume ofinduction motors used in refrigeration chillers of comparablerefrigeration capacity. The result is a smaller sized motor to meet therequired horsepower for a specific compressor assembly.

The permanent magnet motor 52 of an embodiment of the present inventionis compact, efficient, reliable, and relatively quieter thanconventional motors. As the physical size of the screw compressor 12 isreduced, motor 52 used can be scaled in size to fully realize thebenefits of improved fluid flow paths and compressor element shape andsize. Motor 52 is reduced in volume by approximately 30 percent or more,when compared to conventional existing designs for compressor assembliesthat employ induction motors and have refrigeration capacities in excessof 35-tons. The resulting size reduction of embodiments of the presentinvention provides a greater opportunity for efficiency, reliability,and quiet operation through use of less material and smaller dimensionsthan has been achieved through more conventional practices.

Any bearings employed with motor 52 may be rolling element bearings(REB) or hydrodynamic journal bearings. Such bearings may be oillubricated. Oil-free bearing systems may be employed. A special class ofbearing which is refrigerant lubricated is a foil bearing and anotherbearing type uses REB with ceramic balls. Each bearing type hasadvantages and disadvantages that should be apparent to those of skillin the art. Bearings should be selected to facilitate highly efficientoperation of the screw compressor 12 at reduced speeds for capacitymodulation and to minimize rotor dynamics and vibration associated withreduced speeds. Any bearing type may be employed that is suitable ofsustaining rotational speeds in the range of about 2,000 RPM to about20,000 RPM.

The motor rotor 56 and motor stator 54 end turn losses for the permanentmagnet motor 52 are very low compared to some conventional motors,including induction motors. The motor 52, therefore, may be cooled bymeans of fluid 80 (typically, refrigerant). When fluid 80 is employedfor cooling motor 52, fluid 80 may only need to contact the outsidediameter of the stator 54. Cooling the motor 52 in this way allows forthe elimination of the motor cooling feed ring that is typically used ininduction motor stators. Alternatively, refrigerant may be metered tothe outside surface of the stator 54 and to the end turns of the stator54 to cool the motor 52.

In addition, the torque that is needed from motor 52 comes essentiallyfrom the internal pressure distribution in the rotors 42, which is afunction of rotors 42 geometry and the operating conditions. Thatinternal pressure distribution within the rotors 42 provides the loadagainst which the motor 52 has to work. Employing embodiments of thisinvention without a mechanical unloader results in a theoretical torquethat may be essentially constant over a full range of operatingconditions, and for a given operating condition, a ratio of theoreticalto actual torque on the motor 52 that may be approximately constant,despite decay in the actual torque during operation due to changinglosses and leakage, for example. In contrast, for a given operatingcondition, conventional screw compressors invoking a mechanical unloaderwill have significant torque fluctuations or variations over time.

As illustrated in FIG. 4, a variable speed drive 38 may drive the motor52 and in turn, screw compressor 12. The speed of the motor 52 can becontrolled by varying, for example, the frequency of the electric powerthat is supplied to the motor 52. Use of a permanent magnet motor 52 andvariable speed drive 38 moves some conventional motor losses outside ofthe refrigerant loop. The efficiency of the variable speed drive 38,line input to motor shaft output, preferably can achieve a minimum ofabout 95 percent over the system operating range.

The variable speed drive 38 drives the screw compressor 12 at theoptimum, or near optimum, rotational speed at each capacity over thepreselected screw compressor capacity range for a screw compressor 12 ofa given rated capacity. The variable speed drive 38 may be refrigerantcooled, water cooled or air cooled. As mentioned, similar to cooling ofmotor 52, the variable speed drive 38, or portions thereof, may be byusing a refrigerant circulated within the chiller system 10 or by otherconventional cooling means. How the motor 52 and/or variable speed drive38 are cooled should be understood as dependent on the operational andenvironmental conditions in which the motor 52 and/or variable speeddrive 38 reside in operation.

The variable speed drive 38 typically will comprise an electrical powerconverter comprising a line rectifier and line electrical currentharmonic reducer, power circuits and control circuits (such circuitsfurther comprising all communication and control logic, includingelectronic power switching circuits). Conditions in which the screwcompressor 12 is employed may justify employing more than one variablespeed drive 38 for chiller 10.

The variable speed drive 38 can be configured to receive command signals62 from a controller 60 and to generate a control signal 64. Thevariable speed drive 38 will respond, for example, to signals 62received from a microprocessor (also not shown) associated withcontroller system 60 to increase or decrease the speed of the motor 52by changing the frequency of the current supplied to motor 52.Controller 60 may be configured to receive status signals 82 indicativeof an operating point of the screw compressor, and to generate commandsignals that requests the electric motor system to drive the screwcompressor per a preselected operating parameter. Status signals 82 maydeliver similar or different status information depending, for example,on the intended purpose of the sensor selected. Controller 60 maygenerate command signals 62 per a preselected operating parameter, likea torque profile for screw compressor 12. Control signal 64 can drivethe high energy density motor 52 at a rotational speed substantiallygreater than a synchronous motor rotational speed for the rated screwcompressor capacity and drive the motor 52, and in turn at least onescrew rotor 42, at an optimum peripheral velocity independent of therated screw compressor capacity.

The motor 52 and the variable speed drive 38 have power electronics forlow voltage (less than about 600 volts), 50 Hz and 60 Hz applications.Typically, an AC power source (not shown) will supply multiphase voltageand frequency to the variable speed drive 38. The AC voltage or linevoltage delivered to the variable speed drive 38 will typically havenominal values of 200V, 230V, 380V, 415V, 480V, or 600V at a linefrequency of 50 Hz or 60 Hz depending on the AC power source.

By the use of motor 52 and variable speed drive 38, the speed of motor52 can be varied to match varying system requirements. Speed matchingresults in approximately 30 percent more efficient system operationcompared to a compressor without a variable speed drive 38. By runningcompressor 12 at lower speeds when the load on the chiller is not highor at its maximum, sufficient refrigeration effect can be provided tocool the reduced heat load in a manner which saves energy, makes thechiller 10 more economical from a cost-to-run standpoint, andfacilitates highly efficient chiller 10 operation as compared tochillers which are incapable of such load matching at the rotationalspeeds possible via embodiments of the present invention. For example, arated screw compressor capacity of about 100-tons configured accordingto embodiments of the present invention could be efficiently operableover a preselected screw capacity range of about 75-tons to about125-tons.

Screw compressor 12 can be operated at rotational speeds substantiallyhigher than synchronous motor rotational speeds for a given ratedcapacity of the screw compressor 12. The specific optimum speed for therated screw compressor capacity range is a function of screw compressorcapacity and head pressure. Embodiments of the present inventiondramatically improve the discharge porting of fluid 80 and in turn,allow for screw compressor 12 to be operated at a significantlyincreased rotational speed over the rotational speed that gives the bestperformance for conventionally sized rotors and ports. For example, theselected rotational speed for a rated screw compressor capacity of about100-tons, according to embodiments of the present invention, is about5800 revolutions per minute, when the fluid is an R-134a refrigerant. Incontrast, a conventional screw compressor with a rated capacity of about100-tons has a synchronous motor rotational speed is about 3400revolutions per minute, when the fluid is an R-134a refrigerant.

The allowable range of rotational speed for a particular rated capacityof a screw compressor 12 is selected to achieve an optimum peripheralvelocity of at least one of the screw rotors independent of the ratedcapacity of screw compressor 12 that results in a relatively uniformhigh efficiency across the screw compressor product family (e.g.60-tons, 80-tons, 100-tons and 150-tons.) The optimum peripheralvelocity is a constant product of the rotational speed and the radius ofat least one of the rotors 42, typically, the male rotor 42 b. Theapproximately constant optimum peripheral velocity is, for example, inthe range between about 131 feet per second (about 40 meters per second)to about 164 feet per second (about 50 meters per second). In oneembodiment, the approximately constant optimum peripheral velocity isbetween about 42 meters per second (about 137 feet per second) to about45 meters per second (about 147 feet per second) in high pressureapplications, when R-134a refrigerant is the fluid 80. Persons of skillin the art would understand that, for a low pressure application or fora different primary fluid 80, or both, the optimum peripheral velocitymay be different.

The rotational speed of the motor 52 may be selected in combination withconfiguring rotors 42, suction port 76 and discharge port 78 for eachtarget capacity to achieve an approximately constant optimum peripheralvelocity of at least one of the screw rotors 42 regardless of the ratedcapacity of the screw compressor 12. That is, specific combinations ofscrew rotors 42, inlet port 76, discharge port 78 and the operationalrotational speed are selected such that each specific combinationenables each screw compressor 12 to run at approximately the sameoptimum peripheral velocity for each different rated capacity and, inturn, to produce relatively the same high efficiency between or amongeach different rated capacity of screw compressor 12.

Embodiments of the present invention include a method of sizing of atleast two screw compressors 12 with different rated capacities thatachieve approximately constant efficiency across the screw compressorproduct family (e.g. 60-tons, 80-tons, 100-tons and 150-tons.). Byemploying embodiments of this invention, the isoentropic efficiencyversus capacity (in tons) of screw compressor 12 is significantlyincreased, on the order of 15 percent, over a conventional screwcompressor. In addition, because screw compressor 12 is operated atrelatively higher speed, the screw compressor 12 can slowed down on theorder of 20-30 percent of the speed for the operating capacity and stillhave an approximately constant peak efficiency or efficiency plateau ascompared to the efficiency at the rated screw compressor capacity.

The target capacity for each screw compressor 12, each having adifferent rated capacity, is selected. The rotational speed is alsoselected based on the target capacity of each screw compressor 12 tooperate at least one screw rotor 42 in each screw compressor 12 at anapproximately constant optimum peripheral velocity that is independentof the rated capacity of each screw compressor 12. The suction port 76,the at least two screw rotors 42 and the discharge port 78 areconfigured together with the rotational speed selected for each screwcompressor 12.

Specifically, driving screw compressor 12 at an optimum peripheralvelocity allows for each rotor 42 to have a geometry and a profile thatmay remain the same for a wide range of preselected screw compressorcapacities for the rated screw compressor capacity. Each of the rotors42, though, may have a different geometry and a profile for eachdifferent rated screw compressor capacity that will enable at least onescrew rotor to be operated at a selected rotational speed that producesan approximately constant optimum peripheral velocity between or amongeach rated capacity of each screw compressor 12. The volumetric ratio ofthe screw compressor 12 is selected as a function of the loadingconditions in which the screw compressor 12 will be used. By way ofexample, in embodiments of the present invention, more than twovolumetric ratios, potentially four, five or more, are contemplated overa range of rated screw compressor capacities. The volumetric ratio mayalso be such that the system compression ratio and the internalcompression ratio closely match. The rotor 42 profile may be a balanceof the length of the sealing line, flow cross sectional area andblow-hole area size.

The geometry and profile are generally defined, in part, by the numberof lobes in each rotor, the wrap angle, the length of the rotors and thediameter of the rotors, for example. Screw rotor 42 has a profile takenin a plane transverse to the parallel axes of the male rotor 42 b andthe female rotor 42 a. The profile of rotors 42 can be symmetric orasymmetric, and circular, elliptical, parabolic, hyperbolic, forexample. Rack generation of rotors 42 profile may be employed. Selectinga profile of rotors 42 is a balance of the internal leakage path offluid 80 during operation of screw compressor 12 and the portingconfiguration of suction port 76 and discharge port 78, such that screwcompressor 12 has an approximately constant optimum peripheral velocity.

More specifically, for example, at an about 44 m/s optimum peripheralvelocity for at least one rotor 42 of a 100-ton screw compressor, theresulting male rotor 42 b has a wrap angle of about 347 degrees and thefemale rotor 42 a has a wrap angle that is 6/7ths of the male rotor 42b. The wrap angle of the female rotor 42 a varies with the ratio ofnumber of lobes. The female rotor 42 a has a radius of about 2.5 inches(6.35 centimeters) and 7 lobes and the male rotor 42 b has a radius ofabout 3 inches (7.62 centimeters) and 6 lobes. The length of rotors 42is significantly smaller, on the order of about 20-30 percent smaller,than a conventionally sized screw compressor at the rated screwcompressor capacity. A person of skill in the art will appreciate thatanalytical techniques can be employed for other combinations of rotor 42profiles for a given rated screw compressor capacity within the scope ofthe present invention.

Employing a geometry/profile of rotors 42 for a screw compressor 12having a preselected screw compressor range and operable at anapproximately constant optimum peripheral velocity, allows for operationof the screw compressor 12 at 25 or more percent less than the ratedscrew compressor capacity without significant adverse rotor dynamiceffects. Screw compressor 12 has an improved rotor profile thatmaximizes internal flow area, internal friction due to relative motionof the rotor 42 surfaces is minimized, and leakage paths are reduced.This reduced leakage and higher flow tend to increase the screwcompressor 12 efficiency and reduce power wasted, which increasesoverall efficiency.

Referring now to FIG. 4, further details regarding an embodiment of thechiller 10 are presented. In particular, chiller 10 may include acontroller or controller system 60. Controller 60 may be arranged tocommunicate with the variable frequency drive 38, screw compressor 12,condenser 14 and evaporator 20. Chiller 10 may further include one ormore sensors. Sensors 66, 68, 70, 72 and 74, for example, may beemployed to sense and/or communicate torque, suction pressure and/ortemperature, discharge pressure and/or temperature, and/or othermeasurable parameter. Other sensors could be employed depending on theapplication in which screw compressor 12 is used. Signals 82 may becommunicated via wiring, fiber optics, wireless and/or a combination ofwiring, fiber optics and wireless. The sensors 66, 68, 70, 72 and 74communicate status signals 82 to controller 60 with data that areindicative of the operation of various components of the chiller 10.

The controller 60 may include processors, microcontrollers, analogcircuitry, digital circuitry, firmware, and/or software (not shown) thatcooperate to ultimately control operation of the screw compressor 12.The memory may comprise non-volatile memory devices such as flash memorydevices, read only memory (ROM) devices, electricallyerasable/programmable ROM devices, and/or battery backed random accessmemory (RAM) devices to store an array of performance relatedcharacteristics for the screw compressor 12. The memory may furtherinclude instructions which the controller 60 may execute in order tocontrol the operation of the screw compressor 12.

The controller 60 may receive status signals from one or more sensors66, 68, 70, 72 and 74 that provide information regarding operation ofthe screw compressor 12. Based upon the status signals, the controller60 may determine an operating mode and/or operating point of the screwcompressor 12 and may generate, based upon the determined operating modeand/or operating point, one or more command signals 62 to adjust theoperation of the screw compressor 12. The controller 60 may thengenerate command signals 62 that request the motor 52 to operateaccording to a preselected operating parameter(s) (e.g. a torqueprofile). For example, the controller 60 may enable operation at anoptimal torque and speed of screw compressor 12 to minimize losses,mechanical wear and losses.

Further disclosure of a controller system 60 suitable for use withembodiments of the present invention may be found in application U.S.patent Ser. No. 12/544,582 (now U.S. Pat. No. 8,365,544), assigned tothe assignee of the instant application, which is hereby incorporated byreference.

It should be apparent that variations on the control system 60 describedabove will be apparent to those skilled in the art. The control system60 may be implemented with electronic digital, analog, or a combinationof digital/analog control elements and low-voltage wiring. Otherconventional pneumatic tubing, transmitters, controllers, and relays arecontemplated.

In addition, it also will be readily apparent to one of ordinary skillin the art that the compressor system disclosed can be readilyimplemented in other contexts at varying scales. Use of various motortypes, drive mechanisms, and configurations with embodiments of thisinvention should be readily apparent to those of ordinary skill in theart.

Employing embodiments of the present invention, as compared toconventional approaches, increase full-load efficiency, yield higherpart-load efficiency and have a practically constant efficiency over agiven capacity range, controlled independently of power supply frequencyor voltage changes. Also, an advantage of embodiments of the presentinvention is that screw compressors 12 of different rated capacity caneach have a variable capacity and still have the approximately same thelevel of efficiency and without mechanical unloading.

Additional advantages include a reduction in the physical size of thescrew compressor and chiller system arrangement, improved scalability ofthe screw compressors throughout the operating range and a reduction intotal sound levels. Employing embodiments of screw compressor 12 canalso effectively reduce costs for the manufacturer, because it allowsfor one screw compressor at a rated screw compressor capacity (e.g.100-tons) to serve as an efficient screw compressor at a range ofpreselected screw compressor capacity range (e.g. 80 tons and 125 tons)without the need for multiple other screw compressors to be manufacturedat each additional target capacity within the preselected screwcompressor rated capacity range. Practically, embodiments of the presentinvention also allow for lower physical part count and inventory for aproduct family with no loss in capacity or performance due to powersupply because, for a given rated capacity of screw compressor (e.g.100-tons), the screw compressor 12 at 50 Hertz and 60 Hertz are nearlyidentical.

The patentable scope of the invention is defined by the claims asdescribed by the above description. While particular features,embodiments, and applications of the present invention have been shownand described, including the best mode, other features, embodiments orapplications may be understood by one of ordinary skill in the art toalso be within the scope of this invention. It is therefore contemplatedthat the claims will cover such other features, embodiments orapplications and incorporates those features which come within thespirit and scope of the invention.

We claim:
 1. A variable capacity screw compressor comprising: a rotorhousing comprising: a suction port, a working chamber, a discharge port,and at least two screw rotors that comprise a female screw rotor and amale screw rotor positioned within the working chamber for cooperativelycompressing a fluid; the suction port, the at least two screw rotors,and the discharge port are sized in relation to a selected rotationalspeed, the selected rotational speed having an optimum peripheralvelocity of at least one screw rotor of the at least two screw rotorsthat is independent of a peripheral velocity of the at least one screwrotor of the at least two screw rotors at a synchronous motor rotationalspeed for a rated screw compressor capacity, wherein a physical sizingof the at least two screw rotors includes selecting a geometry andprofile of the at least one screw rotor of the at least two screwrotors, the geometry and profile including one or more of a number oflobes, a wrap angle, a length, and a diameter, the optimum peripheralvelocity is a constant product including the selected rotational speedand a radius of the at least one screw rotor of the at least two screwrotors; a permanent magnet motor configured to drive at least one screwrotor of the at least two screw rotors at the selected rotational speedwhen the compressor is at a full-load capacity, the selected rotationalspeed being greater than the synchronous motor rotational speed at therated screw compressor capacity; and a variable speed drive configuredto receive a command signal from a controller and to generate a controlsignal that drives the permanent magnet motor at the selected rotationalspeed, wherein: the suction port is sized and located so as to provide areduced approach velocity and to reduce turbulence for the fluidentering the working chamber, the discharge port is sized larger thantheoretically necessary to provide a thermodynamic optimum size, and thesuction port and the discharge port are configured to reduce a flowloss.
 2. The variable capacity screw compressor of claim 1, wherein theoptimum peripheral velocity is between 141 feet per second to 164 feetper second, where the fluid is R-134A.
 3. The variable capacity screwcompressor of claim 1, wherein the rated screw compressor capacity is100-tons, the fluid is an R-134a refrigerant and the selected rotationalspeed is 5800 revolutions per minute.
 4. The variable capacity screwcompressor of claim 1, wherein the male screw rotor has a wrap angle of347 degrees, the female screw rotor has a radius of 2.5 inches and 7lobes, and the male screw rotor has a radius of 3 inches and 6 lobes. 5.The variable capacity screw compressor of claim 1, wherein the permanentmagnet motor is comprised of high energy density magnetic materials ofat least 20 Mega Gauss Oersted.
 6. The variable capacity screwcompressor of claim 1, wherein the permanent magnet motor provides aconstant ratio of theoretical torque to actual torque for a constantpressure at the discharge port.
 7. The variable capacity screwcompressor of claim 1, wherein the suction port, the at least two screwrotors, and the discharge port are configured in relation to theselected rotational speed so as to provide a constant adiabaticefficiency over a preselected screw compressor rated capacity range. 8.The variable capacity screw compressor of claim 1, wherein the fluid isa low global warming potential (GWP) refrigerant.
 9. The variablecapacity screw compressor of claim 1, wherein the fluid is a refrigerantselected from R-123, R-134a, R-1234yf, R-410A, R-22 or R-32 in gas ormultiple phases.
 10. A refrigeration chiller having at least onerefrigeration circuit comprising: a variable capacity screw compressor,the variable capacity compressor comprises: a rotor housing, a motorhousing and a variable speed drive; the rotor housing further comprises:a suction port, a working chamber, a discharge port, and at least twoscrew rotors that comprise a female screw rotor and a male screw rotorpositioned within the working chamber for cooperatively compressing afluid; the suction port, the at least two screw rotors, and thedischarge port are sized in relation to a selected rotational speed, theselected rotational speed having an optimum peripheral velocity of atleast one screw rotor of the at least two screw rotors that isindependent of a peripheral velocity of the at least one screw rotor ofthe at least two screw rotors at a synchronous motor rotational speedfor a rated screw compressor capacity, wherein a physical sizing of theat least two screw rotors includes selecting a geometry and profile ofthe at least one screw rotor of the at least two screw rotors, thegeometry and profile including one or more of a number of lobes, a wrapangle, a length, and a diameter, the optimum peripheral velocity is aconstant product including the selected rotational speed and a radius ofthe at least one screw rotor of the at least two screw rotors; the motorhousing further comprises: a permanent magnet motor, the permanentmagnet motor is configured to drive at least one screw rotor of the atleast two screw rotors at the selected rotational speed when thecompressor is at a full-load capacity, the selected rotational speedbeing greater than the synchronous motor rotational speed at the ratedscrew compressor capacity; the variable speed drive configured toreceive a command signal from a controller and to generate a controlsignal that drives the permanent magnet motor at the selected rotationalspeed such that the at least one screw rotor operates at the optimumperipheral velocity; wherein: the suction port is sized and located soas to provide a reduced approach velocity and to reduce turbulence forfluid entering the working chamber, the discharge port is sized largerthan theoretically necessary to provide a thermodynamic optimum size,and the suction port and the discharge port are configured to reduce aflow loss; a condenser coupled to the discharge port of the variablecapacity screw compressor, the condenser is configured to cool andcondense fluid received from the discharge port; an expansion valvecoupled to the condenser, the expansion valve is configured to evaporateat least a portion of fluid received from the condenser by loweringpressure of fluid received from the condenser; and an evaporator coupledto the expansion valve, the evaporator is configured to evaporate fluidreceived from the expansion valve and to provide fluid to the suctionport of the variable capacity screw compressor.
 11. The refrigerationchiller of claim 10, wherein the constant optimum peripheral velocity isbetween 141 feet per second to about 164 feet per second, where thefluid is R-134A.
 12. The refrigeration chiller of claim 10, wherein therated screw compressor capacity is 100-tons, the fluid is an R-134arefrigerant and the selected rotational speed is 5800 revolutions perminute for the rated screw compressor capacity.
 13. The refrigerationchiller of claim 12, wherein the male rotor has a wrap angle of 347degrees, the female screw rotor has a radius of 2.5 inches and 7 lobes,and the male screw rotor has a radius of 3 inches and 6 lobes.
 14. Thevariable capacity screw compressor of claim 10, wherein the fluid is alow global warming potential (GWP) refrigerant.
 15. The refrigerationchiller of claim 10, wherein the fluid is a refrigerant selected fromR-123, R-134a, R-1234yf, R-410A, R-22 or R-32 in gas or multiple phases.16. The refrigeration chiller of claim 10, further comprising at leastone economizer cooperatively arranged with one or more of the variablecapacity screw compressor, condenser, or evaporator.
 17. Therefrigeration chiller of claim 10, the controller is further configuredto receive status signals indicative of an operating point of thevariable capacity screw compressor, and to generate the command signalthat requests the permanent magnet motor to drive the variable capacityscrew compressor per a preselected operating parameter.
 18. Therefrigeration chiller of claim 10, wherein the permanent magnet motor isconfigured to receive a variable frequency control signal and to drivethe at least two screw rotors per a received variable frequency controlsignal; the variable frequency drive is further configured to receivethe command signals and is to generate the variable frequency controlsignal to drive the permanent magnet motor per a preselected operatingparameter.